Hydraulic valve play compensating elements, also referred to for short below as HVA, serve, in valve drives of internal combustion engines, for the autonomous compensation of valve play which is generated by production tolerances, thermal expansion and wear in the mechanical transmission chain between cams, cam followers and gas exchange valves. The valve play compensation is generally realized by means of a piston which is guided in an axially moveable manner, and is elastically supported with a piston spring, in a piston housing associated with a valve drive element of the valve drive. The piston spring generates a preload which acts in the transmission chain and which compensates any occurring valve play, with the lifting movement which is to be transmitted to the gas exchange valves being controlled by means of a hydraulically actuable control valve between the piston and piston housing. As a result of the autonomous play compensation, the noise and wear behavior of the valve drive is significantly improved over valve drives without valve play compensation or with manually adjustable valve play compensation. In addition, valve play compensation elements can be utilized as switchable and/or adjustable actuating means in the case of variable control of the valve lifts and/or of the opening and closing times of the intake-side and exhaust-side gas exchange valves, as is described for example in DE 103 00 724 A1.
In a standard design, conventional hydraulic valve play compensation elements use a control valve which is embodied as a non-return valve and which has a control valve closing element and a control valve spring, with the latter acting on the control valve element. The control valve closing element is conventionally embodied as a spherical or sphere-like closing body which is loaded with a force in the closing direction by a coil pressure spring. In this way, the control valve is usually closed, so that the compensating element acts predominantly as a “rigid” actuating element, in which an axial force loading is transmitted directly to the associated gas exchange valve. With said design, however, as a result of the construction, and in particular when the engine is cold and in the event of production inaccuracies in the base circle of the cam, a so-called surging of the control valve with a negative valve play cannot be excluded, which can lead to increased wear and, in the worst case, to engine damage.
Said disadvantages are avoided in the so-called reverse spring hydraulic valve play compensating elements which are also referred to for short below as RSHVA and are known for example from DE 10 2004 018 457 A1. Here, the control valve springs are arranged inversely, so that the control valve ball is acted on in the opening direction, usually delimited by a stop face of a valve cap, and the control valve acts predominantly as a “soft” element. In the case of an RSHVA, at the beginning of a cam lift, an axial force is exerted on the valve play compensating element. Here, a control oil flow is set up from a high-pressure space of an outer housing of the RSHVA to a low-pressure space within a piston, which is movable axially with respect to the housing, of the RSHVA via a piston bore which connects the pressure spaces, which control oil flow acts on the control valve ball. Here, the play compensating element begins to collapse and initially generates an idle stroke until the hydrodynamic and hydrostatic forces of the control oil press the ball counter to the opening spring force and into its valve seat and thereby close the valve, and the gas exchange valve is subsequently actuated.
The influence of the idle stroke, which is typical of the RSHVA, on the overlap of the valve opening and closing times between intake and exhaust valves, and the engine-speed-dependent profile of said influence, can be incorporated in the valve controller in a targeted fashion in order to improve the idle stroke behavior, the thermodynamic efficiency and the pollutant emission optimization of the internal combustion engine.
A problem of the reverse spring valve play compensating elements of the conventional design, having a control valve ball, an opening control valve spring which is embodied as a coil pressure spring which is positioned in an axial piston bore, and a valve cap which holds the control valve ball, has proven to be the relatively complex flow conditions of the control oil between the low-pressure space and the high-pressure space. Tests have shown that in particular the windings of the coil pressure spring hinder the hydraulic flow and make the latter harder to calculate. The valve ball which is acted on with a force by the spring covers the central longitudinal opening of the spring. Since the oil flow also cannot pass the narrow spring windings of the extremely small components, or can pass the narrow spring windings of the extremely small components only to an insufficient extent, the oil flow is restricted substantially to the narrow region between the spring outer periphery and the walls of the piston opening. In addition, the flow through the openings of the conventional valve cap which functions as a stop and guide means for the control valve ball, are made more complicated. Even slight oil-type-related and/or temperature-dependent changes in viscosity of the control oil and production tolerances on the decisive components can therefore lead, in said play compensating element types, to relatively great functional fluctuations during operation, in particular to high idle stroke tolerances, which can unfavorably influence the operating behavior of the internal combustion engine. On the other hand, high demands on the production and material tolerances are expensive and difficult to adhere to.
JP 611 856 07 A discloses a hydraulic valve play compensating element in a hydraulic tappet of a valve drive, in which a disk-shaped spring element which is recessed in the opening direction is arranged below a piston head. Held below the spring element, in a cylindrical valve cap, is a freely moveable control valve ball (a so-called “freeball”). The has a central opening whose diameter is smaller than the diameter of the valve ball, so that in the unpressurized state, the valve ball cannot close off a piston bore. Consequently, in the event of a hydraulic loading of the valve ball in the closing direction, the valve ball is initially pressed against the border of the central opening of the spring disk, and the spring disk is placed in contact with the piston head by overcoming its preload. Here, the control oil which flows past the valve ball at the sides can pass the piston bore via a further opening in the border region of the spring disk until the valve ball closes off the piston bore.
As a result of said arrangement, an engine-speed-dependent oil flow via the control valve is set up in the event of a cam loading by means of a camshaft. With rising cam speed, the oil flow is reduced until the closure of the control valve, as a result of which a piston collapsing movement is reduced and a valve lift of a gas exchange valve is correspondingly enlarged, which leads to more effective cylinder charging in the associated cylinder and therefore to increased engine power. At low engine speeds, in contrast, the valve lift is reduced as a result of the comparatively large idle stroke, as a result of which a fuel saving can be obtained. Here, the spring disk which is embodied as a plate spring acts as a control element.
Said document has the aim of an engine-speed-dependent valve lift variation by means of the idle stroke of an RSHVA, but makes no reference to the flow conditions of the control oil within the RSHVA during its lifting movements. Here, although no spring windings block the flow shortly upstream of or within the piston bore, a disadvantage of the known arrangement is that, in the event of a pressure build-up in the high-pressure space, the oil flow passes initially between the walls of the valve cap and the ball surface, past an upper border of the valve cap and via the outer opening of the spring disk in a ball-ring-shaped flow space, which is formed by the plate-shaped geometry of the spring disk, along the piston head underside, and is deflected upstream thereof to the piston bore before passing the piston bore. Said flow path which is sensitive to boundary layer and turbulence effects can therefore, like in the case of a coil spring within the piston bore, result in undesirably high functional fluctuations in operation in the event of relatively small production and/or material tolerances.
Since, during the closing process of the valve, the control valve ball directly drives the plate spring and in doing so engages partially into the central opening of the spring, high demands must be made in terms of production tolerances on the spring geometry and its assembly position and on the ball in order to preclude significant actuating travel fluctuations. In addition, as a result of the plate-shaped geometry of the spring, lateral clamping forces can act on the ball, which lateral clamping forces can adversely affect correct abutment of the closing element against the piston opening.
In addition, in the event of a relaxation movement, which follows the cam loading, of the compensating element in the cam base circle when the gas exchange valve is closed, during which relaxation movement the piston and piston housing are pushed apart again by the piston spring until there is no play present between the cam follower and the cam, the return flow of the control oil from the low-pressure space (reservoir) to the high-pressure space can be unfavorably hindered by means of the relatively restricted flow path via the ball, plate spring and valve cap.